U.S. patent number 9,222,706 [Application Number 13/581,477] was granted by the patent office on 2015-12-29 for refrigeration cycle apparatus and operating method of same.
This patent grant is currently assigned to MITSUBISHI ELECTRIC CORPORATION. The grantee listed for this patent is Takeshi Hatomura, Masayuki Kakuda, Hideaki Nagata, Yusuke Shimazu, Keisuke Takayama. Invention is credited to Takeshi Hatomura, Masayuki Kakuda, Hideaki Nagata, Yusuke Shimazu, Keisuke Takayama.
United States Patent |
9,222,706 |
Takayama , et al. |
December 29, 2015 |
Refrigeration cycle apparatus and operating method of same
Abstract
A refrigeration cycle apparatus achieves efficient operation by
constantly recovering power in a wide operating range. The
refrigeration cycle apparatus regulates a pressure of a high
pressure side by changing either one or both of an opening degree
of the intermediate-pressure bypass valve and an opening degree of
the pre-expansion valve on the basis of a density ratio that is
obtained from an inflow refrigerant density of the expander and an
inflow refrigerant density of the sub-compressor in an actual
operating state and a design volume ratio that has been expected at
the time of design and that is obtained from a stroke volume of the
sub-compressor, a stroke volume of the expander, and a ratio of a
flow rate of the refrigerant flowing to the sub-compressor.
Inventors: |
Takayama; Keisuke (Tokyo,
JP), Shimazu; Yusuke (Tokyo, JP), Kakuda;
Masayuki (Tokyo, JP), Nagata; Hideaki (Tokyo,
JP), Hatomura; Takeshi (Tokyo, JP) |
Applicant: |
Name |
City |
State |
Country |
Type |
Takayama; Keisuke
Shimazu; Yusuke
Kakuda; Masayuki
Nagata; Hideaki
Hatomura; Takeshi |
Tokyo
Tokyo
Tokyo
Tokyo
Tokyo |
N/A
N/A
N/A
N/A
N/A |
JP
JP
JP
JP
JP |
|
|
Assignee: |
MITSUBISHI ELECTRIC CORPORATION
(Chiyoda-Ku, Tokyo, JP)
|
Family
ID: |
44672524 |
Appl.
No.: |
13/581,477 |
Filed: |
March 25, 2010 |
PCT
Filed: |
March 25, 2010 |
PCT No.: |
PCT/JP2010/002121 |
371(c)(1),(2),(4) Date: |
August 28, 2012 |
PCT
Pub. No.: |
WO2011/117924 |
PCT
Pub. Date: |
September 29, 2011 |
Prior Publication Data
|
|
|
|
Document
Identifier |
Publication Date |
|
US 20120318001 A1 |
Dec 20, 2012 |
|
Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F25B
1/10 (20130101); F25B 49/027 (20130101); F25B
2400/14 (20130101); F25B 2600/0261 (20130101); F25B
41/39 (20210101); F25B 2309/061 (20130101); F25B
2600/2513 (20130101) |
Current International
Class: |
F25B
1/10 (20060101); F25B 9/00 (20060101); F25B
49/02 (20060101); F25B 49/00 (20060101); F25B
13/00 (20060101) |
Field of
Search: |
;62/286.2,498,401,402,324.6,87,513,510,196 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
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|
2886450 |
|
Apr 2007 |
|
CN |
|
101506597 |
|
Aug 2009 |
|
CN |
|
58-188557 |
|
Dec 1983 |
|
JP |
|
2004-108683 |
|
Apr 2004 |
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JP |
|
2004-150748 |
|
May 2004 |
|
JP |
|
3708536 |
|
Oct 2005 |
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JP |
|
2006242491 |
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Sep 2006 |
|
JP |
|
2008-39237 |
|
Feb 2008 |
|
JP |
|
2009-162438 |
|
Jul 2009 |
|
JP |
|
10-2007-0046974 |
|
May 2007 |
|
KR |
|
WO 2009/047898 |
|
Apr 2009 |
|
WO |
|
2009/142014 |
|
Nov 2009 |
|
WO |
|
Other References
Office Action issued on Mar. 13, 2014, by the Chinese Patent Office
in corresponding Chinese Patent Application No. 201080065731.4 and
an English translation of the Office Action. (6 pages). cited by
applicant .
Japanese Office Action (Notification of Reasons for Rejection)
dated Jun. 4, 2013, issued in corresponding Japanese Patent
Application No. JP2012-506667, and English Translation of the
Office Action. (5 pgs.). cited by applicant .
International Search Report (PCT/ISA/210) issued on Jun. 29, 2010,
by the Japanese Patent Office as the International Searching
Authority for International Application No. PCT/JP2010/002121.
cited by applicant.
|
Primary Examiner: Elve; M. Alexandra
Assistant Examiner: Crenshaw; Henry
Attorney, Agent or Firm: Buchanan Ingersoll & Rooney
PC
Claims
The invention claimed is:
1. A refrigeration cycle apparatus, comprising: a main compressor
configured to compress a refrigerant; a radiator configured to
radiate heat of the refrigerant compressed by the main compressor;
an expander configured to reduce a pressure of the refrigerant that
has passed through the radiator; an evaporator configured to
evaporate the refrigerant that has been depressurized by the
expander; a sub-compressor having a discharge side connected to a
compression chamber of the main compressor, the sub-compressor
using power, which is generated in the expander when reducing the
pressure of the refrigerant, to compress a portion of the
refrigerant passing through the evaporator to an intermediate
pressure; an intermediate-pressure bypass configured to connect a
refrigerant outflow side of the sub-compressor and a refrigerant
inflow side of the main compressor to each other; an
intermediate-pressure bypass valve arranged in the
intermediate-pressure bypass, the intermediate-pressure bypass
valve configured to control a flow rate of the refrigerant flowing
through the intermediate-pressure bypass; a pre-expansion valve
arranged between a refrigerant outflow side of the radiator and a
refrigerant inflow side of the expander, the pre-expansion valve
configured to reduce the pressure of the refrigerant flowing into
the expander; and a controller configured to control an operation
of the intermediate-pressure bypass valve and an operation of the
pre-expansion valve, wherein the controller configured to regulate
a pressure of a high pressure side by changing either one or both
of an opening degree of the intermediate-pressure bypass valve and
an opening degree of the pre-expansion valve.
2. The refrigeration cycle apparatus of claim 1, wherein the
controller is configured to increase the pressure of the high
pressure side by changing either one or both of the opening degree
of the intermediate-pressure bypass valve and the opening degree of
the pre-expansion valve when a density ratio that is obtained from
an inflow refrigerant density of the expander and an inflow
refrigerant density of the sub-compressor in an actual operating
state is larger than a design volume ratio that is obtained from a
stroke volume of the sub-compressor, a stroke volume of the
expander, and a ratio of a flow rate of the refrigerant flowing to
the sub-compressor, and the controller is configured to decrease
the pressure of the high pressure side by changing either one or
both of the opening degree of the intermediate-pressure bypass
valve and the opening degree of the pre-expansion valve when the
density ratio in the actual operating state is smaller than the
design volume ratio.
3. The refrigeration cycle apparatus of claim 1, wherein the
controller is configured to regulate the pressure of the high
pressure side based on a comparative result of a target discharge
temperature and a discharge temperature that is detected at a
refrigerant outflow side of the main compressor.
4. The refrigeration cycle apparatus of claim 1, wherein the
controller is configured to regulate the pressure of the high
pressure side based on a comparative result of a target degree of
superheat and a degree of superheat of the refrigerant flowing out
from the evaporator.
5. The refrigeration cycle apparatus of claim 1, wherein the
controller is configured to regulate the pressure of the high
pressure side based on a comparative result of a target degree of
supercooling and a degree of supercooling of the refrigerant
flowing out from the radiator.
6. The refrigeration cycle apparatus of claim 1, wherein the
controller is configured to regulate the pressure of the high
pressure side by operating the pre-expansion valve when the opening
degree of the intermediate-pressure bypass valve is at a minimum
opening degree, and by operating the intermediate-pressure bypass
valve when the opening degree of the pre-expansion valve is at a
maximum opening degree.
7. The refrigeration cycle apparatus of claim 1, wherein the main
compressor is a two stage compressor, and the refrigerant
discharged from the sub-compressor is injected to a passage
connecting a low-stage-side compression chamber and a
latter-stage-side compression chamber to each other.
8. The refrigeration cycle apparatus of claim 1, wherein a
refrigerant that enters a supercritical state on the high-pressure
side is used as the refrigerant.
9. An operating method of a refrigeration cycle apparatus,
comprising the steps of: compressing a refrigerant with a main
compressor; radiating heat of the refrigerant compressed by the
main compressor with a radiator; reducing a pressure of the
refrigerant that has passed through the radiator with an expander;
evaporating the refrigerant that has been depressurized by the
expander with an evaporator; using power, which has been generated
in the expander when reducing the pressure of the refrigerant, for
compressing a portion of the refrigerant passing through the
evaporator to an intermediate pressure with a sub-compressor;
injecting the refrigerant compressed to the intermediate pressure
by the sub-compressor to a midway position of a compression process
of the main compressor; connecting a refrigerant outflow side of
the sub-compressor and a refrigerant inflow side of the main
compressor to each other with an intermediate-pressure bypass;
controlling a flow rate of the refrigerant flowing through the
intermediate-pressure bypass with an intermediate-pressure bypass
valve; reducing the pressure of the refrigerant that is flowing
between a refrigerant outflow side of the radiator and a
refrigerant inflow side of the expander and that is flowing into
the expander with a pre-expansion valve; and regulating a pressure
of a high pressure side by changing either one or both of an
opening degree of the intermediate-pressure bypass valve and an
opening degree of the pre-expansion valve on the basis of a density
ratio that is obtained from an inflow refrigerant density of the
expander and an inflow refrigerant density of the sub-compressor in
an actual operating state and a design volume ratio and that is
obtained from a stroke volume of the sub-compressor, a stroke
volume of the expander, and a ratio of a flow rate of the
refrigerant flowing to the sub-compressor.
10. The operating method of the refrigeration cycle apparatus of
claim 9, wherein the pressure of the high pressure side is
increased by changing either one or both of the opening degree of
the intermediate-pressure bypass valve and the opening degree of
the pre-expansion valve when the density ratio in the actual
operating state is larger than the design volume ratio.
11. The operating method of the refrigeration cycle apparatus of
claim 9, wherein the pressure of the high pressure side is reduced
by changing either one or both of the opening degree of the
intermediate-pressure bypass valve and the opening degree of the
pre-expansion valve when the density ratio in the actual operating
state is smaller than the design volume ratio.
Description
TECHNICAL FIELD
The present invention relates to refrigeration cycle apparatuses
and operating methods of the same, and more particularly relates to
a refrigeration cycle apparatus and an operating method of the same
that uses a refrigerant that undergoes transition to a
supercritical state, that includes coaxially coupled compressor and
expander, that recovers expansion power that is generated when the
refrigerant is expanded, and that uses the expansion power for
compressing the refrigerant.
BACKGROUND ART
In recent years, there has been focus on refrigeration cycle
apparatuses that uses, as its refrigerant, carbon dioxide
(hereinafter, referred to as CO.sub.2), which has zero ozone
depleting potential and a markedly small global warming potential
as compared with those of chlorofluorocarbons. The critical
temperature of the CO.sub.2 refrigerant is as low as 31.06 degrees
C. When a temperature higher than this temperature is used, the
refrigerant at a high-pressure side (from the outlet of a
compressor, to a radiator, and then to the inlet of a decompressor)
of the refrigeration cycle apparatus enters a supercritical state
in which no condensation occurs, thereby decreasing operating
efficiency (COP) of the refrigeration cycle apparatus as compared
with conventional refrigerants. Hence, means for increasing COP is
important to refrigeration cycle apparatuses using a CO.sub.2
refrigerant.
As such means, there is suggested a refrigeration cycle that is
provided with an expander instead of a decompressor and that
recovers pressure energy during expansion as power. Meanwhile, in a
refrigeration cycle apparatus with a configuration in which a
positive displacement compressor and an expander are coupled with a
single shaft, when VC is a stroke volume of the compressor and VE
is a stroke volume of the expander, a ratio of the volumetric
circulation rate of the refrigerants respectively flowing through
the compressor and the expander is determined by VC/VE (a design
volume ratio). When DC is density of the refrigerant at an outlet
of an evaporator (the refrigerant flowing into the compressor) and
DE is density of the refrigerant at an outlet of an radiator (the
refrigerant flowing into the expander), a relationship of
"VC.times.DC=VE.times.DE," that is, a relationship of "VC/VE=DE/DC"
is established since the mass circulation rate of the refrigerants
respectively flowing through the compressor and the expander are
equivalent. VC/VE (the design volume ratio) is a constant that is
determined at the time of design of the device. The refrigeration
cycle tends to balance itself so that DE/DC (the density ratio) is
always constant (hereinafter, this is called "constraint of
constant density ratio").
However, use conditions of the refrigeration cycle apparatus are
not necessarily constant, and hence if the design volume ratio
expected at the time of the design differs to the density ratio in
the actual operating state, it would be difficult to regulate the
pressure of the high pressure side to an optimal pressure due to
the "constraint of constant density ratio."
Owing to this, there is suggested a configuration and a control
method for regulating the pressure of the high pressure side to the
optimal pressure by providing a bypass that bypasses the expander
and controlling the amount of refrigerant flowing into the expander
(for example, see Patent Literature 1).
Also, there is suggested a configuration and a control method for
regulating the pressure of the high pressure side to the optimal
pressure by providing a compression bypass that bypasses a phase
from a midway position of a compression process of a main
compressor to completion of the compression process, by providing a
sub-compressor in the compression bypass, and by controlling the
amount of refrigerant flowing into the sub-compressor (for example,
see Patent Literature 2).
CITATION LIST
Patent Literature
Patent Literature 1: Japanese Patent No. 3708536 (Claim 1, FIG. 1,
etc.)
Patent Literature 2: Japanese Unexamined Patent Application
Publication No. 2009-162438 (Claim 1, FIG. 1, etc.)
SUMMARY OF INVENTION
Technical Problem
Patent Literature 1 describes the configuration and the control
method that can regulate the pressure of the high pressure side to
the optimal pressure by distributing the refrigerant to the bypass
that bypasses the expander if the density ratio in the actual
operating state is smaller than the design volume ratio; however,
the refrigerant flowing through a bypass valve may be subjected to
isenthalpic change because of expansion loss. Hence, there is a
problem in which an effect of increasing refrigerating effect
obtained by the isentropic change while the expander recovers the
expansion energy decreases.
In addition, if the amount of refrigerant that bypasses the
expander is large, the rotation speed of the expander becomes low
and a lubrication state of a sliding portion is degraded. If the
rotation speed of the expander becomes excessively low, problems
arise such as stagnation of oil in a passage of the expander that
causes degradation of reliability such as exhaustion of the oil in
the compressor and starting up with the stagnated refrigerant at
the time of restart.
Further, Patent Literature 2 attempts to address the
above-described problems by not bypassing the expander. However,
since the bypass valve is provided at the inlet of the
sub-compressor, the pressure at the inlet of the sub-compressor
decreases due to pressure loss, and compression power increases by
that amount. There is a problem in that the increasing effect of
the operating efficiency is decreased.
The present invention is made to address the above problems, and an
object of the invention is to provide a refrigeration cycle
apparatus and an operating method capable of providing highly
efficient operation by constantly recovering power in a wide
operating range even if it is difficult to regulate the pressure of
the high pressure side to the optimal pressure due to the
constraint of constant density ratio.
Solution to Problem
A refrigeration cycle apparatus according to the invention includes
a main compressor compressing a refrigerant; a radiator radiating
heat of the refrigerant compressed by the main compressor; an
expander reducing a pressure of the refrigerant that has passed
through the radiator; an evaporator evaporating the refrigerant
which has been depressurized by the expander; a sub-compressor
having a discharge side connected to a midway position of a
compression process of the main compressor, the sub-compressor
using power, which is generated in the expander when reducing the
pressure of the refrigerant, to compress a portion of the
refrigerant passing through the evaporator to an intermediate
pressure; an intermediate-pressure bypass connecting a refrigerant
outflow side of the sub-compressor and a refrigerant inflow side of
the main compressor to each other; an intermediate-pressure bypass
valve being provided in the intermediate-pressure bypass, the
intermediate-pressure bypass valve controlling a flow rate of the
refrigerant flowing through the intermediate-pressure bypass; a
pre-expansion valve being provided between a refrigerant outflow
side of the radiator and a refrigerant inflow side of the expander,
the pre-expansion valve reducing the pressure of the refrigerant
flowing into the expander; and a controller controlling an
operation of the intermediate-pressure bypass valve and an
operation of the pre-expansion valve. The controller regulates a
pressure of a high pressure side by changing either one or both of
an opening degree of the intermediate-pressure bypass valve and an
opening degree of the pre-expansion valve on the basis of a density
ratio that is obtained from an inflow refrigerant density of the
expander and an inflow refrigerant density of the sub-compressor in
an actual operating state and a design volume ratio that has been
expected at the time of design and that is obtained from a stroke
volume of the sub-compressor, a stroke volume of the expander, and
a ratio of a flow rate of the refrigerant flowing to the
sub-compressor.
An operating method of a refrigeration cycle apparatus according to
the invention includes the steps of: compressing a refrigerant with
a main compressor; radiating heat of the refrigerant compressed by
the main compressor with a radiator; reducing a pressure of the
refrigerant that has passed through the radiator with an expander;
evaporating the refrigerant which has been depressurized by the
expander with an evaporator; using power, which has been generated
in the expander when reducing the pressure of the refrigerant, for
compressing a portion of the refrigerant passing through the
evaporator to an intermediate pressure with a sub-compressor;
injecting the refrigerant compressed to the intermediate pressure
by the sub-compressor to a midway position of a compression process
of the main compressor; connecting a refrigerant outflow side of
the sub-compressor and a refrigerant inflow side of the main
compressor to each other with an intermediate-pressure bypass;
controlling a flow rate of the refrigerant flowing through the
intermediate-pressure bypass with an intermediate-pressure bypass
valve; reducing the pressure of the refrigerant that is flowing
between a refrigerant outflow side of the radiator and a
refrigerant inflow side of the expander and that is flowing into
the expander with a pre-expansion valve; and regulating a pressure
of a high pressure side by changing either one or both of an
opening degree of the intermediate-pressure bypass valve and an
opening degree of the pre-expansion valve on the basis of a density
ratio that is obtained from an inflow refrigerant density of the
expander and an inflow refrigerant density of the sub-compressor in
an actual operating state and a design volume ratio that has been
expected at the time of design and that is obtained from a stroke
volume of the sub-compressor, a stroke volume of the expander, and
a ratio of a flow rate of the refrigerant flowing to the
sub-compressor.
Advantageous Effects of Invention
With the refrigeration cycle apparatus and the operating method of
the refrigeration cycle according to the invention, highly
efficient operation can be achieved by recovering the power in the
wide operating range and regulating the pressure of the high
pressure side through the control of the intermediate-pressure
bypass valve and the pre-expansion valve even if it is difficult to
regulate the pressure of the high pressure side to the optimal
pressure due to the constraint of constant density ratio.
BRIEF DESCRIPTION OF DRAWINGS
FIG. 1 is a circuit configuration diagram schematically showing a
refrigerant circuit configuration of a refrigeration cycle
apparatus according to an Embodiment of the invention.
FIG. 2 is a schematic longitudinal section showing a sectional
configuration of a main compressor.
FIG. 3 is a P-h diagram showing transition of a refrigerant during
a cooling operation of the refrigeration cycle apparatus according
to an Embodiment of the invention.
Fig. 4 is a P-h diagram showing transition of the refrigerant
during a heating operation of the refrigeration cycle apparatus
according to an Embodiment of the invention.
FIG. 5 is a flowchart showing the flow of control processing
executed by a controller.
FIG. 6 is an explanatory view showing an operation during
cooperative control of an intermediate-pressure bypass valve and a
pre-expansion valve.
FIG. 7 is a P-h diagram showing transition of the refrigerant when
an operation of closing the pre-expansion valve 6 is performed
during the cooling operation executed by the refrigeration cycle
apparatus according to an Embodiment of the invention.
FIG. 8 is a P-h diagram showing transition of the refrigerant when
an operation of opening the intermediate-pressure bypass valve is
performed during the cooling operation executed by the
refrigeration cycle apparatus according to an Embodiment of the
invention.
FIG. 9 is a P-h diagram showing a part of transition of a carbon
dioxide refrigerant.
FIG. 10 is a circuit configuration diagram schematically showing a
refrigerant circuit configuration of a refrigeration cycle
apparatus according to an Embodiment of the invention.
DESCRIPTION OF EMBODIMENTS
Embodiment of the invention will be described below with reference
to the drawings.
FIG. 1 is a circuit configuration diagram schematically showing a
refrigerant circuit configuration of a refrigeration cycle
apparatus 100 according to Embodiment of the invention. FIG. 2 is a
schematic longitudinal section showing a sectional configuration of
a main compressor 1. FIG. 3 is a P-h diagram showing transition of
a refrigerant during a cooling operation of the refrigeration cycle
apparatus 100. FIG. 4 is a P-h diagram showing transition of the
refrigerant during a heating operation of the refrigeration cycle
apparatus 100. FIG. 5 is a flowchart showing the flow of control
processing executed by a controller 83. FIG. 6 is an explanatory
view showing an operation during cooperative control of an
intermediate-pressure bypass valve 9 and a pre-expansion valve 6. A
circuit configuration and an operation of the refrigeration cycle
apparatus 100 will be described with reference to FIGS. 1 to 6.
The refrigeration cycle apparatus 100 according to Embodiment is
used in devices equipped with a refrigeration cycle that circulates
a refrigerant and is used, for example, in a refrigerator, a
freezer, a vending machine, an air-conditioning apparatus (for
domestic use, industrial use, or vehicles, for example), a
refrigeration apparatus, or a water heater. It should be noted that
the dimensional relationships of components in FIG. 1 and other
subsequent drawings may be different from the actual ones. In
addition, in FIG. 1 and other subsequent drawings, components
applied with the same reference signs correspond to the same or
equivalent components. This is common through the full text of the
description. Further, forms of components described in the full
text of the description are mere examples, and the components are
not limited to the described forms of components.
The refrigeration cycle apparatus 100 can constantly recover power
in a wide operating range and can perform efficient operations. In
particular, the advantageous effect is large when a carbon dioxide
refrigerant in which a high-pressure side enters a supercritical
state is used.
The refrigeration cycle apparatus 100 at least includes the main
compressor 1, an outdoor heat exchanger 4, an expander 7, an indoor
heat exchanger 21, and a sub-compressor 2. Also, the refrigeration
cycle apparatus 100 includes a first four-way valve 3 serving as a
refrigerant passage switching unit, a second four-way valve 5
serving as a refrigerant passage switching unit, the pre-expansion
valve 6, an accumulator 8, an intermediate-pressure bypass valve 9,
and a check valve 10. Further, the refrigeration cycle apparatus
100 includes a controller 83 that controls the overall control of
the refrigeration cycle apparatus 100.
The main compressor 1 compresses a refrigerant, which is sucked by
an electric motor 102 and a shaft 103 driven by the electric motor
102, and turns the refrigerant into a high-temperature
high-pressure state. This main compressor 1 may be constituted by,
for example, a capacity-controllable inverter compressor. It is to
be noted that the details of the main compressor 1 is described
later with reference to FIG. 2.
The outdoor heat exchanger 4 functions as a radiator in which the
refrigerant therein radiates heat during a cooling operation, and
functions as an evaporator in which the refrigerant therein
evaporates during a heating operation. For example, the outdoor
heat exchanger 4 exchanges heat between the air, which is supplied
from a fan (not shown), and the refrigerant.
The outdoor heat exchanger 4 has a heat transfer pipe, through
which the refrigerant passes, and a fin for increasing a heat
transferring area between the refrigerant flowing through the heat
transfer pipe and the outdoor air. The outdoor heat exchanger 4 is
configured to exchange heat between the refrigerant and the air
(the outdoor air). The outdoor heat exchanger 4 functions as an
evaporator during the heating operation. The outdoor heat exchanger
4 evaporates and gasifies (vaporizes) the refrigerant. On the other
hand, the outdoor heat exchanger 4 functions as a condenser or a
gas cooler (hereinafter, referred to as a condenser) during the
cooling operation. In some cases, the outdoor heat exchanger 4 may
not completely gasify or vaporize the refrigerant, and may turn the
refrigerant into a two-phase mixture of gas and liquid (two-phase
gas-liquid refrigerant).
The indoor heat exchanger 21 functions as an evaporator in which
the refrigerant therein evaporates during the cooling operation,
and functions as a radiator in which the refrigerant therein
radiates heat during the heating operation. The indoor heat
exchanger 21 exchanges heat between the air, which is supplied from
a fan (not shown), and the refrigerant.
The indoor heat exchanger 21 has a heat transfer pipe, through
which the refrigerant passes, and a fin for increasing a heat
transferring area between the refrigerant flowing through the heat
transfer pipe and the outdoor air. The indoor heat exchanger 21 is
configured to exchange heat between the refrigerant and the indoor
air. The indoor heat exchanger 21 functions as an evaporator during
the cooling operation. The indoor heat exchanger 21 evaporates the
refrigerant and gasifies (vaporizes) the refrigerant. On the other
hand, the indoor heat exchanger 21 functions as a condenser or a
gas cooler (hereinafter, referred to as condenser) during the
heating operation.
The expander 7 reduces the pressure of the refrigerant passing
therethrough. Power that is generated when the pressure of the
refrigerant is reduced is transferred to the sub-compressor 2
through a driving shaft 43. The sub-compressor 2 is connected to
the expander 7 through the driving shaft 43. The sub-compressor 2
is driven by power that is generated when the expander 7 reduces
the pressure of the refrigerant, and the sub-compressor 2
compresses the refrigerant. The sub-compressor 2 is connected in
parallel to the main compressor 1 in a lower-pressure side.
Regarding the expander 7 and the sub-compressor 2, the driving
shaft 43 recovers expansion power that is generated when the
expander 7 expands (reduces the pressure of) the refrigerant and
the sub-compressor 2 uses the recovered expansion power and
compresses the refrigerant. The expander 7 and the sub-compressor 2
are of a positive displacement type, and employ a form of, for
example, scroll type. The sub-compressor 2 and the expander 7 are
housed in a hermetically sealed container 84. The sub-compressor 2
is connected to the expander 7 through the driving shaft 43, so
that the driving shaft 43 recovers the power that is generated in
the expander 7 and transfers the power to the sub-compressor 2.
Thus, the refrigerant is also compressed in the sub-compressor
2.
The first four-way valve 3 is provided in a discharge piping 35 of
the main compressor 1, and has a function of switching the flow
direction of the refrigerant in accordance with an operating mode.
By switching the first four-way valve 3, connection is made between
the outdoor heat exchanger 4 and the main compressor 1, between the
indoor heat exchanger 21 and the accumulator 8, between the indoor
heat exchanger 21 and the main compressor 1, or between the outdoor
heat exchanger 4 and the accumulator 8. That is, the first four-way
valve 3 performs switching in accordance with the operating mode
relating to cooling and heating based on an instruction of the
controller 83, and hence switches the passage of the
refrigerant.
The second four-way valve 5 connects the expander 7 to the outdoor
heat exchanger 4 or the indoor heat exchanger 21 in accordance with
the operating mode. By switching the second four-way valve 5,
connection is made between the outdoor heat exchanger 4 and the
pre-expansion valve 6, between the indoor heat exchanger 21 and the
expander 7, between the indoor heat exchanger 21 and the
pre-expansion valve 6, or between the outdoor heat exchanger 4 and
the expander 7. That is, the second four-way valve 5 performs
switching in accordance with the operating mode relating to cooling
and heating based on an instruction of the controller 83, and hence
switches the passage of the refrigerant.
During the cooling operation, the first four-way valve 3 is
switched such that the refrigerant flows from the main compressor 1
to the outdoor heat exchanger 4 and flows from the indoor heat
exchanger 21 to the accumulator 8, and the second four-way valve 5
is switched such that the refrigerant flows from the outdoor heat
exchanger 4 to the indoor heat exchanger 21 through the
pre-expansion valve 6 and the expander 7. In contrast, during the
heating operation, the first four-way valve 3 is switched such that
the refrigerant flows from the main compressor 1 to the indoor heat
exchanger 21 and flows from the outdoor heat exchanger 4 to the
accumulator 8, and the second four-way valve 5 is switched such
that the refrigerant flows from the indoor heat exchanger 21 to the
outdoor heat exchanger 4 through the pre-expansion valve 6 and the
expander 7. With the second four-way valve 5, the direction of the
refrigerant passing through the expander 7 is the same in either of
the cooling operation and the heating operation.
The pre-expansion valve 6 is provided upstream of the expander 7
and expands the refrigerant by reducing the pressure of the
refrigerant, and may be one having a variably controllable opening
degree such as an electronic expansion valve. To be more specific,
the pre-expansion valve 6 is provided in a refrigerant passage 34
arranged between the second four-way valve 5 and the inlet of the
expander 7 (i.e., the refrigerant outflow side of the radiator (the
outdoor heat exchanger 4 or the indoor heat exchanger 21) and the
refrigerant inflow side of the expander 7), and regulates the
pressure of the refrigerant flowing into the expander 7.
The accumulator 8 is provided on the suction side of the main
compressor 1 and has a function of retaining the liquid refrigerant
so as to prevent the liquid from returning to the main compressor 1
when a failure has occurred in the refrigeration cycle apparatus
100 or during a transient response of the operating state due to a
change in operation control. That is, the accumulator 8 has a
function of retaining excessive refrigerant in the refrigerant
circuit of the refrigeration cycle apparatus 100 and preventing the
main compressor 1 from being damaged when the liquid refrigerant
returns to the main compressor 1 and the sub-compressor 2 by a
large amount.
The intermediate-pressure bypass valve 9 is provided in an
intermediate-pressure bypass piping (an intermediate-pressure
bypass) 33 that causes the refrigerant to bypass from a discharge
piping 31 of the sub-compressor 2 to a suction piping 32 of the
main compressor 1, and controls the flow rate of the refrigerant
flowing through the intermediate-pressure bypass piping 33. The
intermediate-pressure bypass valve 9 may be one having a variably
controllable opening degree such as an electronic expansion valve.
By adjusting the opening degree of the intermediate-pressure bypass
valve 9, the intermediate pressure, which is the discharge pressure
of the sub-compressor 2, can be regulated.
The check valve 10 is provided in the discharge piping 31 of the
sub-compressor 2, and adjusts the flow direction of the refrigerant
flowing into the main compressor 1 to one direction (a direction
from the sub-compressor 2 to the main compressor 1). By providing
this check valve 10, backflow of the refrigerant occurring when the
discharge pressure of the sub-compressor 2 becomes lower than the
pressure of a compression chamber 108 of the main compressor 1 can
be prevented.
The controller 83 controls the driving frequency of the main
compressor 1, the rotation speeds of the fans (not shown) provided
near the outdoor heat exchanger 4 and the indoor heat exchanger 21,
switching of the first four-way valve 3, switching of the second
four-way valve 5, the opening degree of the expander 7, the opening
degree of the pre-expansion valve 6, the opening degree of the
intermediate-pressure bypass valve 9, and the like.
It is to be noted that description is given in Embodiment assuming
that the refrigeration cycle apparatus 100 uses carbon dioxide
(CO.sub.2) as its refrigerant. Carbon dioxide has characteristics
such as zero ozone depleting potential and a small global warming
potential as compared with conventional chlorofluorocarbon based
refrigerants. However, the refrigerant is not limited to carbon
dioxide, and other single refrigerants, mixed refrigerants (for
example, a mixed refrigerant of carbon dioxide and diethyl ether),
or the like that undergoes transition to a supercritical state may
be used as the refrigerant.
In the refrigeration cycle apparatus 100, the main compressor 1,
the sub-compressor 2, the first four-way valve 3, the second
four-way valve 5, the outdoor heat exchanger 4, the pre-expansion
valve 6, the expander 7, the accumulator 8, the
intermediate-pressure bypass valve 9, and the check valve 10 are
housed in an outdoor unit 81. In addition, in the refrigeration
cycle apparatus 100, the controller 83 is also housed in the
outdoor unit 81. Further, in the refrigeration cycle apparatus 100,
the indoor heat exchanger 21 is housed in an indoor unit 82. FIG. 1
exemplarily illustrates a state in which the single outdoor unit 81
(the outdoor heat exchanger 4) is connected to the single indoor
unit 82 (the indoor heat exchanger 21) through a liquid pipe 36 and
a gas pipe 37; however, the numbers of connected outdoor units 81
and indoor units 82 are not particularly limited.
In addition, temperature sensors (a temperature sensor 51, a
temperature sensor 52, and a temperature sensor 53) are provided in
the refrigeration cycle apparatus 100. Temperature information
detected by these temperature sensors is sent to the controller 83,
and used for control of components of the refrigeration cycle
apparatus 100.
The temperature sensor 51 is provided in the discharge piping 35 of
the main compressor 1, detects the discharge temperature of the
main compressor 1, and may be constituted by, for example, a
thermistor. The temperature sensor 52 is provided near the outdoor
heat exchanger 4 (for example, on the outer surface), detects the
temperature of the air flowing into the outdoor heat exchanger 4,
and may be constituted by, for example, a thermistor. The
temperature sensor 53 is provided near the indoor heat exchanger 21
(for example, on the outer surface), detects the temperature of the
air flowing into the indoor heat exchanger 21, and may be
constituted by, for example, a thermistor.
It is to be noted that the installation positions of the
temperature sensor 51, the temperature sensor 52, and the
temperature sensor 53 are not limited to the positions shown in
FIG. 1. For example, the temperature sensor 51 may be installed at
any position where the temperature of the refrigerant discharged
from the main compressor 1 can be detected, the temperature sensor
52 may be installed at any position where the temperature of the
air flowing into the outdoor heat exchanger 4 can be detected, and
the temperature sensor 53 may be installed at any position where
the temperature of the air flowing into the indoor heat exchanger
21 can be detected.
The configuration and operation of the main compressor 1 will be
described with reference to FIG. 2. The main compressor 1 is
configured such that a shell 101 which forms the outline of the
main compressor 1 houses therein the electric motor 102 serving as
a driving source, the shaft 103 serving as the driving shaft
rotationally driven by the electric motor 102, an oscillating
scroll 104 attached to a distal end of the shaft 103 and
rotationally driven together with the shaft 103, a fixed scroll 105
arranged above the oscillating scroll 104 and having a spiral body
that meshes with a spiral body of the oscillating scroll 104, and
the like. Also, an inflow piping 106 that is connected to the
suction piping 32, an outflow piping 112 that is connected to the
discharge piping 35, and an injection piping 114 that is connected
to the discharge piping 31 are connected to the shell 101.
A low-pressure space 107 that is in communication with the inflow
piping 106 is formed in the shell 101, at an outermost periphery
portion of the spiral bodies of the oscillating scroll 104 and the
fixed scroll 105. A high-pressure space 111 that is in
communication with the outflow piping 112 is formed in an upper
inner portion of the shell 101. A plurality of compression
chambers, whose capacities relatively change, are formed between
the spiral body of the oscillating scroll 104 and the spiral body
of the fixed scroll (for example, a compression chamber 108 and a
compression chamber 109 shown in FIG. 1). The compression chamber
109 illustrates a compression chamber formed at substantially
center portions of the oscillating scroll 104 and the fixed scroll
105. The compression chamber 108 illustrates a compression chamber
formed during midway of a compression process, at the outside of
the compression chamber 109.
An outflow port 110 that allows the compression chamber 109 to be
in communication with the high-pressure space 111 is provided at
the substantially center portion of the fixed scroll 105. An
injection port 113 that allows the compression chamber 108 to be in
communication with the injection piping 114 is provided at the
midway position of the compression process of the fixed scroll 105.
In addition, an Oldham ring (not shown) for stopping rotation
movement of the oscillating scroll 104 during eccentric turning
movement is arranged in the shell 101. This Oldham ring provides
the function of stopping the rotation movement and a function of
allowing orbital motion of the oscillating scroll 104.
It is to be noted that the fixed scroll 105 is fixed inside the
shell 101. In addition, the oscillating scroll 104 performs orbital
motion without performing the rotation movement relative to the
fixed scroll 105. Further, the electric motor 102 includes at least
a stator that is fixed inside the shell 101, and a rotor that is
arranged so as to be rotatable inside an inner peripheral surface
of the stator and that is fixed to the shaft 103. The stator has a
function of rotatably driving the rotor when the stator is
energized. The rotor has a function of being rotatably driven and
rotating the shaft 103 when the stator is energized.
The operation of the main compressor 1 will be briefly described.
When the electric motor 102 is energized, a torque is generated at
the stator and the rotor constituting the electric motor 102, and
the shaft 103 is rotated. Since the oscillating scroll 104 is
mounted at the distal end of the shaft 103, the oscillating scroll
104 performs the orbital motion. The compression chamber moves
toward the center while the capacity of the compression chamber is
decreased by the turning movement of the oscillating scroll 104,
and hence the refrigerant is compressed.
The refrigerant compressed in the sub-compressor 2 and discharged
therefrom passes through the discharge piping 31 and the check
valve 10. This refrigerant then flows from the injection piping 114
into the main compressor 1. Meanwhile, the refrigerant passing
through the suction piping 32 flows from the inflow piping 106 into
the main compressor 1. The refrigerant that has flowed in from the
inflow piping 106 flows into the low-pressure space 107, is
enclosed in the compression chamber, and is gradually compressed.
Then, when the compression chamber reaches the compression chamber
108 at the midway position of the compression process, the
refrigerant flows from the injection port 113 into the compression
chamber 108.
That is, the refrigerant that has flowed in from the injection
piping 114 is mixed with the refrigerant that has flowed in from
the inflow piping 106 in the compression chamber 108. Then, the
mixed refrigerant is gradually compressed and reaches the
compression chamber 109. The refrigerant that has reached the
compression chamber 109 passes through the outflow port 110 and the
high-pressure space 111, is discharged outside the shell 101
through the outflow piping 112, and passes through the discharge
piping 35.
The operating action of the refrigeration cycle apparatus 100 will
be described.
<Cooling Operation Mode>
The operation executed by the refrigeration cycle apparatus 100
during the cooling operation will be described with reference to
FIGS. 1 and 3. It is to be noted that signs A to G shown in FIG. 1
correspond to signs A to G shown in FIG. 3. In addition, in the
cooling operation mode, the first four-way valve 3 and the second
four-way valve 5 are controlled in a state indicated by "solid
lines" in FIG. 1. Here, the highs and lows of the pressure in the
refrigerant circuit and the like of the refrigeration cycle
apparatus 100 are not determined in relation to a reference
pressure, but relative pressures as the result of an increase in
pressure by the main compressor 1 or the sub-compressor 2 and a
reduction in pressure by the pre-expansion valve 6 or the expander
7 are respectively expressed as a high pressure and a low pressure.
In addition, the highs and lows of the temperature are similarly
expressed.
During the cooling operation, first, a low-pressure refrigerant is
sucked into the main compressor 1 and the sub-compressor 2. The
low-pressure refrigerant sucked into the sub-compressor 2 is
compressed by the sub-compressor 2 and turns into an
intermediate-pressure refrigerant (from a state A to a state B).
The intermediate-pressure refrigerant that has been compressed by
the sub-compressor 2 is discharged from the sub-compressor 2, and
is introduced into the main compressor 1 through the discharge
piping 31 and the injection piping 114. The intermediate-pressure
refrigerant is mixed with the refrigerant sucked into the main
compressor 1, is further compressed by the main compressor 1, and
turns into a high-temperature high-pressure refrigerant (from the
state B to a state C). The high-temperature high-pressure
refrigerant that has been compressed by the main compressor 1 is
discharged from the main compressor 1, passes through the first
four-way valve 3, and flows into the outdoor heat exchanger 4.
The refrigerant that has flowed into the outdoor heat exchanger 4
radiates heat by exchanging heat with the outdoor air supplied to
the outdoor heat exchanger 4, transfers heat to the outdoor air,
and turns into a low-temperature high-pressure refrigerant (from
the state C to a state D). The low-temperature high-pressure
refrigerant flows out from the outdoor heat exchanger 4, passes
through the second four-way valve 5, and passes through the
pre-expansion valve 6. The pressure of the low-temperature
high-pressure refrigerant is reduced when passing through the
pre-expansion valve 6 (from the state D to a state E). The
refrigerant whose pressure has been reduced by the pre-expansion
valve 6 is sucked into the expander 7. The pressure of the
refrigerant that has been sucked into the expander 7 is reduced and
the refrigerant becomes low in temperature. Hence, the refrigerant
turns into a refrigerant with low quality (from the state E to a
state F).
At this time, power is generated in the expander 7 as the result of
the reduction in pressure of the refrigerant. The power is
recovered by the driving shaft 43, is transferred to the
sub-compressor 2, and is used for the compression of the
refrigerant by the sub-compressor 2. The refrigerant whose pressure
has been reduced by the expander 7 is discharged from the expander
7, passes through the second four-way valve 5, and then flows out
from the outdoor unit 81. The refrigerant flowing out from the
outdoor unit 81 flows through the liquid pipe 36 and flows into the
indoor unit 82.
The refrigerant that has flowed into the indoor unit 82 flows into
the indoor heat exchanger 21, receives heat from the indoor air
supplied to the indoor heat exchanger 21 and evaporates, and turns
into a refrigerant still low in pressure but with high quality
(from the state F to a state G). Accordingly, the indoor air is
cooled. This refrigerant flows out from the indoor heat exchanger
21, also flows out from the indoor unit 82, flows through the gas
pipe 37, and flows into the outdoor unit 81. The refrigerant that
has flowed into the outdoor unit 81 passes through the first
four-way valve 3, flows into the accumulator 8, and is sucked into
the main compressor 1 and the sub-compressor 2 again.
The refrigeration cycle apparatus 100 repeats the above-described
operation and, accordingly, the heat of the indoor air is
transferred to the outdoor air; hence, the indoor air is
cooled.
<Heating Operation Mode>
The operation executed by the refrigeration cycle apparatus 100
during the heating operation will be described with reference to
FIGS. 1 and 4. It is to be noted that signs A to G shown in FIG. 1
correspond to signs A to G shown in FIG. 4. In addition, in the
heating operation mode, the first four-way valve 3 and the second
four-way valve 5 are controlled in a state indicated by "broken
lines" in FIG. 1. Here, the highs and lows of the pressure in the
refrigerant circuit and the like of the refrigeration cycle
apparatus 100 are not determined in relation to a reference
pressure, but relative pressures as the result of an increase in
pressure by the main compressor 1 or the sub-compressor 2 and a
reduction in pressure by the pre-expansion valve 6 or the expander
7 are respectively expressed as a high pressure and a low pressure.
In addition, the highs and lows of the temperature are similarly
expressed.
During the heating operation, first, a low-pressure refrigerant is
sucked into the main compressor 1 and the sub-compressor 2. The
low-pressure refrigerant sucked into the sub-compressor 2 is
compressed by the sub-compressor 2 and turns into an
intermediate-pressure refrigerant (from a state A to a state B).
The intermediate-pressure refrigerant that has been compressed by
the sub-compressor 2 is discharged from the sub-compressor 2, and
is introduced into the main compressor 1 through the discharge
piping 31 and the injection piping 114. The intermediate-pressure
refrigerant is mixed with the refrigerant sucked into the main
compressor 1, is further compressed by the main compressor 1, and
turns into a high-temperature high-pressure refrigerant (from the
state B to a state G). The high-temperature high-pressure
refrigerant that has been compressed by the main compressor 1 is
discharged from the main compressor 1, passes through the first
four-way valve 3, and flows out from the outdoor unit 81.
The refrigerant that has flowed out from the outdoor unit 81 flows
through the gas pipe 37 and flows into the indoor unit 82. The
refrigerant that has flowed into the indoor unit 82 flows into the
indoor heat exchanger 21, radiates heat by exchanging heat with the
indoor air supplied to the indoor heat exchanger 21, transfers heat
to the indoor air, and turns into a low-temperature high-pressure
refrigerant (from the state G to the state F). Accordingly, the
indoor air is heated. This low-temperature high-pressure
refrigerant flows out from the indoor heat exchanger 21, flows out
from the indoor unit 82, flows through the liquid pipe 36, and
flows into the outdoor unit 81. The refrigerant that has flowed
into the outdoor unit 81 passes through the second four-way valve
5, and passes through the pre-expansion valve 6. The pressure of
the low-temperature high-pressure refrigerant is reduced when
passing through the pre-expansion valve 6 (from the state F to a
state E).
The refrigerant whose pressure has been reduced by the
pre-expansion valve 6 is sucked into the expander 7. The pressure
of the refrigerant that has been sucked into the expander 7 is
reduced and the refrigerant becomes low in temperature. Hence, the
refrigerant turns into a refrigerant with low quality (from the
state E to a state D). At this time, power is generated in the
expander 7 as the result of the reduction in pressure of the
refrigerant. The power is recovered by the driving shaft 43, is
transferred to the sub-compressor 2, and is used for the
compression of the refrigerant by the sub-compressor 2. The
refrigerant whose pressure has been reduced by the expander 7 is
discharged from the expander 7, passes through the second four-way
valve 5, and then flows into the outdoor heat exchanger 4. The
refrigerant that has flowed into the outdoor heat exchanger 4
receives heat from the outdoor air supplied to the outdoor heat
exchanger 4 and evaporates, and turns into a refrigerant still low
in pressure but with high quality (from the state D to a state
C).
The refrigerant flows out from the outdoor heat exchanger 4, passes
through the first four-way valve 3, flows into the accumulator 8,
and is sucked into the main compressor 1 and the sub-compressor 2
again.
The refrigeration cycle apparatus 100 repeats the above-described
operation and, accordingly, the heat of the outdoor air is
transferred to the indoor air; hence, the indoor air is heated.
Here, the flow rates of the refrigerant of the sub-compressor 2 and
the expander 7 will be described.
It is assumed that GE is a flow rate of the refrigerant flowing
through the expander 7, and GC is a flow rate of the refrigerant
flowing through the sub-compressor 2. Also, when it is assumed that
W is a ratio of the flow rate (referred to as diversion ratio) of
the refrigerant flowing to the sub-compressor 2 among the total
flow rate of the refrigerant flowing to the main compressor 1 and
the sub-compressor 2, the relationship between GE and GC is
expressed by Expression (1) as follows. GC=W.times.GE Expression
(1)
Hence, when VC is a stroke volume of the sub-compressor 2, VE is a
stroke volume of the expander 7, DC is an inflow refrigerant
density of the sub-compressor 2, and DE is an inflow refrigerant
density of the expander 7, the constraint of constant density ratio
is expressed by Expression (2) as follows. VC/VE/W=DE/DC Expression
(2)
In addition, the diversion ratio W may be determined such that the
recovery power of the expander 7 and the compression power of the
sub-compressor 2 are substantially equivalent to each other. To be
more specific, when hE is a specific enthalpy of the inlet of the
expander 7, hF is a specific enthalpy of the outlet of the expander
7, hA is a specific enthalpy of the inlet of the sub-compressor 2,
and hB is a specific enthalpy of the outlet of the sub-compressor
2, the diversion ratio W may be determined to satisfy Expression
(3) as follows. hE-hF=W.times.(hB-hA) Expression (3)
Since the refrigeration cycle apparatus 100 injects the refrigerant
to the main compressor 1 after the sub-compressor 2 compresses part
of the low-pressure refrigerant to the intermediate pressure, an
electric input of the main compressor 1 can be reduced by the
amount of the compression power of the sub-compressor 2.
Next, the cooling operation when a density ratio (DE/DC) in an
actual operating state differs to a design volume ratio (VC/VE/W)
expected at the time of design will be described.
<Cooling Operation when (DE/DC)>(VC/VE/W)>
A cooling operation when the density ratio (DE/DC) in the actual
operating state is larger than the design volume ratio (VC/VE/W)
expected at the time of the design will be described. In this case,
due to the constraint of constant density ratio, the refrigeration
cycle tends to balance itself so that the inlet refrigerant density
(DE) decreases while the pressure of the high pressure side is kept
in a low pressure state. However, when in the state in which the
pressure of the high pressure side is lower than a desirable
pressure, operating efficiency decreases.
Owing to this, if the intermediate-pressure bypass valve 9 is not
in a totally closed state, the intermediate-pressure bypass valve 9
is operated in the closing direction, so as to increase the
intermediate pressure and increase the required compression power
of the sub-compressor 2. Then, the expander 7 will tend to decrease
its rotation speed; hence, the refrigeration cycle will tend to
balance itself towards increasing the inlet density of the expander
7.
Alternatively, if the intermediate-pressure bypass valve 9 is in a
totally closed state, the pre-expansion valve 6 is operated in the
closing direction, so as to expand the refrigerant flowing into the
expander 7 (from the state D to a state E2) as shown in FIG. 7 and
decrease the refrigerant density. Then, the refrigeration cycle
will tend to balance itself towards increasing the inlet density of
the expander 7. FIG. 7 is a P-h diagram showing transition of the
refrigerant when an operation of closing the pre-expansion valve 6
is performed during the cooling operation executed by the
refrigeration cycle device 100.
To be more specific, in the cooling operation when
(DE/DC)>(VC/VE/W), the refrigeration cycle apparatus 100
controls the intermediate-pressure bypass valve 9 to be closed or
the pre-expansion valve 6 to be closed so that the refrigeration
cycle is balanced towards increasing the pressure of the high
pressure side. Owing to this, the refrigeration cycle apparatus 100
can increase the pressure of the high pressure side and regulate
the pressure of the high pressure side to the desirable pressure.
In addition, since no refrigerant bypasses the expander 7,
efficient operation can be achieved. It is to be noted that the
pressure of the high pressure side refers to a pressure from the
outflow port of the main compressor 1 to the pre-expansion valve 6,
and may be a pressure at any position between the outflow port of
the main compressor 1 and the pre-expansion valve 6.
<Cooling Operation when (DE/DC)<(VC/VE/W)>
A cooling operation when the density ratio (DE/EC) in the actual
operating state is smaller than the design volume ratio (VC/VE/W)
expected at the time of the design will be described. In this case,
due to the constraint of constant density ratio, the refrigeration
cycle tends to balance itself so that the inlet refrigerant density
(DE) increases while the pressure of the high pressure side is kept
in a high pressure state. However, when in the state in which the
pressure of the high pressure side is higher than the desirable
pressure, operating efficiency decreases.
Owing to this, if the pre-expansion valve 6 is not in a fully
opened state, the pre-expansion valve 6 is operated in the opening
direction, so that the refrigerant flowing into the expander 7 does
not expand and the refrigerant density is increased. Then, the
refrigeration cycle will tend to balance itself towards decreasing
the inlet density of the expander 7.
Alternatively, if the pre-expansion valve 6 is in a fully opened
state, the intermediate-pressure bypass valve 9 is operated in the
opening direction. The operation of the refrigerant cycle at this
time will be described with reference to FIG. 8. FIG. 8 is a P-h
diagram showing transition of the refrigerant when an operation of
opening the intermediate-pressure bypass valve 9 is performed
during the cooling operation executed by the refrigeration cycle
device 100.
The sub-compressor 2 compresses the refrigerant flowing out from
the accumulator 8 to an intermediate pressure (from the state G to
the state B). Part of the refrigerant discharged from the
sub-compressor 2 passes through the check valve 10 and is injected
to the main compressor 1. Also, residual part of the refrigerant
discharged from the sub-compressor 2 passes through the
intermediate-pressure bypass valve 9, and joins the refrigerant
flowing through the suction piping 32 of the main compressor 1 (a
state A2). The refrigerant in the state A2 sucked into the main
compressor 1 is mixed with the refrigerant compressed to the
intermediate pressure and injected, and is further compressed (a
state C2). Then, the intermediate-pressure is reduced, the required
compression power of the sub-compressor 2 is decreased, and the
expander 7 tends to increase its rotation speed; hence the
refrigeration cycle tends to balance itself towards decreasing the
inlet density of the expander 7.
To be more specific, in the cooling operation when
(DE/DC)<(VC/VE/W), the refrigeration cycle apparatus 100
controls the pre-expansion valve 6 to be opened or the
intermediate-pressure bypass valve 9 to be opened so that the
refrigeration cycle is balanced towards decreasing the pressure of
the high pressure side. Owing to this, the refrigeration cycle
apparatus 100 can decrease the pressure of the high pressure side
and regulate the pressure of the high pressure side to the
desirable pressure. In addition, since no refrigerant bypasses the
expander 7, efficient operation can be achieved.
<Heating Operation when (DE/DC).noteq.(VC/VE/W)>
There may be a case in which the density ratio (DE/DC) in the
actual operating state of the heating operation differs to the
design volume ratio (VC/VE/W) expected at the time of the design.
The operations of the sub-compressor 2 and the expander 7 are
controlled in a similar manner to that of the cooling operation,
and hence the description is omitted.
Next, as a specific operating method of the intermediate-pressure
bypass valve 9 and the pre-expansion valve 6, the flow of a control
processing executed by the controller 83 will be described with
reference to a flowchart shown in FIG. 5.
The refrigeration cycle apparatus 100 uses the correlation between
the pressure of the high pressure side and the discharge
temperature and executes the control of the intermediate-pressure
bypass valve 9 and the pre-expansion valve 6 based on the discharge
temperature that is relatively inexpensively measured, without
depending on the pressure of the high pressure side that needs an
expensive sensor for measurement.
When the refrigeration cycle apparatus 100 is in operation, the
optimal pressure of the high pressure side is not always constant.
Hence, in the refrigeration cycle apparatus 100, storage means such
as a ROM mounted on the controller 83 stores, in advance, data such
as the outdoor air temperature detected by the temperature sensor
52 and the indoor temperature detected by the temperature sensor
53, in a form of a table. Further, the controller 83 determines a
target discharge temperature from the data stored in the storage
means (step 201). Next, the controller 83 fetches a detection value
(discharge temperature) from the temperature sensor 51 (step 202).
The controller 83 compares the target discharge temperature
determined in step 201 and the discharge temperature fetched in
step 202 (step 203).
If the discharge temperature is lower than the target discharge
temperature (step 203; YES), since the pressure of the high
pressure side tends to be lower than the optimal pressure of the
high pressure side, the controller 83 determines first whether or
not the intermediate-pressure bypass valve 9 is totally closed
(step 204). If the intermediate-pressure bypass valve 9 is totally
closed (step 204; YES), the controller 83 operates the
pre-expansion valve 6 in the closing direction (step 205) to reduce
the pressure of the refrigerant flowing into the expander 7, to
decrease the refrigerant density, and to increase the pressure of
the high pressure side and the discharge temperature. If the
intermediate-pressure bypass valve 9 is not totally closed (step
204; NO), the controller 83 operates the intermediate-pressure
bypass valve 9 in the closing direction (step 206) to increase the
intermediate pressure, to increase the required compression force
of the sub-compressor 2, and to increase the pressure of the high
pressure side and the discharge temperature.
In contrast, if the discharge temperature is higher than the target
discharge temperature (step 203; NO), since the pressure of the
high pressure side tends to be higher than the optimal pressure of
the high pressure side, the controller 83 determines first whether
or not the pre-expansion valve 6 is fully opened (step 207). If the
pre-expansion valve 6 is fully opened (step 207; YES), the
controller 83 operates the intermediate-pressure bypass valve 9 in
the opening direction (step 208) to reduce the intermediate
pressure, to decrease the required compression force of the
sub-compressor 2, and to reduce the pressure of the high pressure
side and the discharge temperature. Also, if the pre-expansion
valve 6 is not fully opened (step 207; NO), the controller 83
operates the pre-expansion valve 6 in the opening direction (step
209) so as not to reduce the pressure of the refrigerant flowing
into the expander 7, and to reduce the pressure of the high
pressure side and the discharge temperature.
After these steps, the control returns to step 201, and repeats
step 201 to step 209. Since such control is executed, the
cooperative control of the intermediate-pressure bypass valve 9 and
the pre-expansion valve 6 can be achieved as shown in FIG. 6. To be
more specific, the controller 83 regulates the pressure of the high
pressure side by operating the pre-expansion valve 6 when the
pressure of the high pressure side is low and the opening degree of
the intermediate-pressure bypass valve is at its minimum, and by
operating the intermediate-pressure bypass valve 9 when the
pressure of the high pressure side is high and the opening degree
of the pre-expansion valve 6 is at its maximum. It is to be noted
that, in FIG. 6, the horizontal axis indicates the high/low level
of the pressure of the high pressure side, the upper section of the
vertical axis indicates the opening degree of the pre-expansion
valve 6, and the lower section of the vertical axis indicates the
opening degree of the intermediate-pressure bypass valve 9.
As described above, the refrigeration cycle apparatus 100 uses the
expander 7 that has difficulty in maintaining the pressure of the
high pressure side to an optimal pressure due to the constraint of
constant density ratio. However, even if the density ratio (DE/DC)
in the actual operating state is smaller or larger than the design
volume ratio (VC/VE/W) expected at the time of design, the pressure
of the high pressure side is regulated to the desirable pressure
and power is reliably recovered without having the refrigerant to
bypass the expander 7 by way of the opening-degree operation of the
intermediate-pressure bypass valve 9 and the pre-expansion valve 6.
Owing to this, the refrigeration cycle apparatus 100 is capable of
achieving an operation that does not drop the operating efficiency
or the operating performance, and can ensure reliability of the
expander 7 and the main compressor 1.
Also, in the refrigeration cycle apparatus 100, the target value of
the opening-degree operation of the intermediate-pressure bypass
valve 9 and the pre-expansion valve 6 is the discharge temperature
of the main compressor 1; however, a pressure sensor may be
provided in the discharge piping 35 of the main compressor 1 and
the target value may be controlled based on the discharge
pressure.
In the refrigeration cycle apparatus 100, the target value of the
opening-degree operation of the intermediate-pressure bypass valve
9 and the pre-expansion valve 6 is the discharge temperature of the
main compressor 1; however, the target value may be a degree of
superheat at the refrigerant outlet of the indoor heat exchanger 21
functioning as an evaporator during the cooling operation. In this
case, the controller 83 may determine the target degree of
superheat on the basis of information from a pressure sensor, which
detects a low-pressure-side pressure, arranged in the refrigerant
piping between the outlet of the expander 7 and the main compressor
1 or the sub-compressor 2 and information from a temperature sensor
that detects a refrigerant outlet temperature of the indoor heat
exchanger 21, in which the information is stored, in advance, in a
ROM or the like in a form of a table.
In addition, the target degree of superheat may be set by providing
a controller in the indoor unit 82. In this case, the target degree
of superheat may be sent to the controller 83 through communication
between the indoor unit 82 and the outdoor unit 81 in a wired or
wireless manner.
Further, regarding the relationship between the pressure of the
high pressure side and the degree of superheat of the evaporator,
it will be such that higher the pressure of the high pressure side,
larger the degree of superheat, and lower the pressure of the high
pressure side, smaller the degree of superheat. Thus, control may
be executed such that the discharge temperature in step 203 in the
flowchart of FIG. 5 is replaced with the degree of superheat.
Also, in the refrigeration cycle apparatus 100, the target value of
the opening-degree operation of the intermediate-pressure bypass
valve 9 and the pre-expansion valve 6 is the discharge temperature
of the main compressor 1; however, the target value may be a degree
of supercooling at the refrigerant outlet of the indoor heat
exchanger 21 functioning as a condenser during the heating
operation.
Embodiment exemplarily shows the case in which CO.sub.2 is used as
the refrigerant of the refrigeration cycle apparatus 100. In a case
in which such a refrigerant is used, when the air temperature of
the condenser is high, the refrigerant is not condensed at the
high-pressure side unlike conventional chlorofluorocarbon based
refrigerants and enters a supercritical cycle. Hence, the degree of
supercooling cannot be calculated from a saturated pressure and a
temperature. Owing to this, as shown in FIG. 9, a pseudo-saturation
pressure and a pseudo-saturation temperature Tc may be determined
based on an enthalpy at the critical point, and the difference with
a refrigerant temperature Tco may be used as a pseudo degree of
supercooling Tsc (see the following Expression (4)). Tsc=Tc-Tco
Expression (4)
Further, regarding the relationship between the pressure of the
high pressure side and the degree of superheat of the condenser, it
will be such that higher the pressure of the high pressure side,
larger the degree of supercooling, and lower the pressure of the
high pressure side, smaller the degree of supercooling. Thus,
control may be executed such that the discharge temperature in step
203 in the flowchart of FIG. 5 is replaced with the degree of
supercooling.
With the refrigeration cycle apparatus 100, phenomena causing
concern when the amount by which the refrigerant bypasses the
expander 7 is large leading to degradation of reliability, such as
degradation in the lubrication state in the sliding portion due to
low rotation speed of the expander 7, exhaustion of oil in the
compressor due to stagnation of oil in the expander and the passage
of the expander 7, and starting up with the stagnated refrigerant
at the time of restart, can be reduced.
With the refrigeration cycle apparatus 100, since an expander
bypass valve is not needed, there will be no expansion loss that is
caused when the refrigerant is expanded by the expander bypass
valve and a decrease in refrigerating effect at the evaporator can
be made small.
With the refrigeration cycle apparatus 100, even when the
sub-compressor 2 can hardly compress the refrigerant, a part of the
circulating refrigerant is made to flow into the sub-compressor 2.
Owing to this, with the refrigeration cycle apparatus 100, as
compared with a case in which the entire amount of circulating
refrigerant is made to flow, the sub-compressor 2 will not degrade
the performance by becoming a passage resistance of the
refrigerant. The case in which the sub-compressor 2 can hardly
compress the refrigerant is, for example, a case in which the
difference between the pressure of the high pressure side and the
low-pressure-side pressure is small and the power recovered by the
expander 7 is excessively small, such as a cooling operation with a
low outdoor air temperature, or a heating operation with a low
indoor temperature.
The refrigeration cycle apparatus 100 is configured such that the
compression function is divided into the main compressor 1 having
the driving source, and the sub-compressor 2 driven by the power of
the expander 7. Hence, with the refrigeration cycle apparatus 100,
the structural design and functional design can be divided. Hence,
problems in view of designing and manufacturing are less than those
of integrated apparatuses of the driving source, expander, and
compressor.
In addition, in the refrigeration cycle apparatus 100, the
refrigerant compressed by the sub-compressor 2 is injected to the
compression chamber 108 of the main-compressor 1. Alternatively,
for example, the compression mechanism of the main compressor 1 may
be a two-stage compression mechanism and the refrigerant may be
injected to a passage connecting a low-stage-side compression
chamber and a latter-stage-side compression chamber. Still
alternatively, the main compressor 1 may be configured to execute
two-stage compression with a plurality of compressors.
In the refrigeration cycle apparatus 100, the outdoor heat
exchanger 4 and the indoor heat exchanger 21 are each a heat
exchanger that exchanges heat with air; however, it is not limited
thereto, and may be a heat exchanger that exchanges heat with other
heat mediums, such as water or brine.
In addition, in the refrigeration cycle apparatus 100, it is
exemplarily described that the refrigerant passage is switched in
accordance with the operation mode relating to cooling and heating,
with the first four-way valve 3 and the second four-way valve 5;
however, it is not limited thereto. For example, the configuration
may be such that a two-way valve, a three-way valve, or a check
valve switches the refrigerant passage.
Reference Signs List
1 main compressor; 2 sub-compressor; 3 first four-way valve; 4
outdoor heat exchanger; 5 second four-way valve; 6 pre-expansion
valve; 7 expander; 8 accumulator; 9 intermediate-pressure bypass
valve; 10 check valve; 21 indoor heat exchanger; 31 discharge
piping; 32 suction piping; 33 intermediate-pressure bypass piping;
34 refrigerant passage; 35 discharge piping; 36 liquid pipe; 37 gas
pipe; 43 driving shaft; 51 temperature sensor; 52 temperature
sensor; 53 temperature sensor; 81 outdoor unit; 82 indoor unit; 83
controller; 84 hermetically sealed container; 100 refrigeration
cycle apparatus; 101 shell; 102 electric motor; 103 shaft; 104
oscillating scroll; 105 fixed scroll; 106 inflow piping; 107
low-pressure space; 108 compression chamber; 109 compression
chamber; 110 outflow port; 111 high-pressure space; 112 outflow
piping; 113 injection port; 114 injection piping.
* * * * *